There are various types of linear electric motors and electromagnetic actuators that can move an object along a limited length linear path and are controlled electronically. These actuators utilize electric energy and convert it into kinetic energy of a moving object when accelerating the object from a stationary position. They also compensate for any motion-related losses and decelerate the object to a stop at the end of travel while dissipating the kinetic energy of the object or converting it partially back to some other form of energy.
Utilization of such electric and electromagnetic actuators for driving of intake and exhaust poppet valves of an internal combustion engine is one of the many applications of these actuator mechanisms. Many such known actuators are designed for controlling the opening and closing of engine valves for implementation of variable valve timing (VVT) and variable valve lift (VVL). This significant attention to VVT and VVL is explained by corresponding performance improvements of a conventional internal combustion engines in terms of improved fuel efficiency and reduced emissions when VVT and VVL are used.
One embodiment of an electronically controlled linear actuator for a valve mechanism involves direct utilization of a linear electric motor for opening or closing the valve during the engine operation. Due to a very short interval of time allowed for the opening and closing events, on the order of few milliseconds, the electric motor that can open or close a conventional engine valve in the required amount of time would be too big and inefficient for practical implementation in the engine. This inefficiency arises mainly due to ohmic losses in the motor coils. In addition, if the motor is required to counter a spring force, there are likewise ohmic losses in countering the spring force in addition to the inertial forces for accelerating the valve.
With reference to FIG. 1, there is shown a prior art electronic valve system with the valve 100 oriented roughly in a half-open position. The valve controls the flow of gas through manifold 102 as part of, for example, an internal combustion engine. The net force of the springs 104 and 106 is roughly zero in this configuration. Computer-controlled electromagnets 108 and 110 act on the armature 112 to form a simple linear motor that works in conjunction with the springs 104 and 106 on the valve stem 114 to cause valve transitions. Valve closure is accomplished when the valve 100 is brought into contact with the valve seat 116. The upper spring 104 is constrained by a stationary retainer 118 and the lower spring 106 is constrained by the stationary cylinder head 120. Hash marks 122 are used to emphasize that a component is stationary for illustrative clarity, but not all stationary components have hash mark in the figures.
FIGS. 2A and 2B show possible generic spring force 200 and spring energy profiles 252 which are generally indicative of the relationship between the valve displacement and the spring restoring force such as that shown in FIG. 1. During operation of the valve 100 of FIG. 1, the upper electromagnet 108 holds the armature 112 in an upper position with the valve 100 closed while a relatively large spring force is pulling the valve 100 downward. To open the valve 100, the upper electromagnet is turned off and the valve 100 is released. The combined mass/spring system of valve 100, valve stem 114, armature 112 and springs 104 and 106 begins a nearly sinusoidal oscillation. Following a half-cycle of motion, the lower electromagnet is activated and latches the valve in the open position. Since the spring force is large in the open position, the acceleration of the valve 100 is largest just as it latched. A similar sequence closes the valve 100. The electromagnets 108 and 110 together with the armature 112 therefore form a kind of linear motor.
Because the electronic engine valve of FIG. 1 has maximum spring forces when the electromagnets 108 and 110 latch the valve 100 open or closed, the valve 100 undergoes an impulse in its 3rd derivative of motion known as the “jerk.” This impulse is associated with vibration and noise even when the valve 100 is latched at zero velocity. In practice, the latching forces can cause the armature 112 to impact the electromagnets 108 and 110 causing additional noise and vibration. Moreover, the valve 100 must contact the valve seat 116 at nearly the same time that the armature 112 contacts the upper electromagnet 108. Otherwise the valve 100 may leak or may impact the valve seat 116 with excessive velocity. This creates a tolerance problem that increases the cost of manufacturing and complicates the control of the electronic engine valve. The problem of positioning the valve seat 116 accurately relative to other components (e.g., the face of upper electromagnet 108) is common and the present discussion refers to it as the “valve seat positioning problem” when it occurs.
The generalized solid curve 252 in FIG. 2B can be seen to specifically illustrate the total spring potential energy in the springs 104 and 106 of FIG. 1. In the middle position, the potential energy is defined to be zero, and as the armature 112 is deflected in either direction from this middle position the spring potential energy grows in a characteristic quadratic form of curve 252 for ideal springs obeying Hooke's law. The motion of the valve 100 under the forces of the springs 104 and 106 is analogous to the motion of ball 250 rolling in the energy well defined by the total potential energy curve 252 of the springs 104 and 106. As may be clearly understood from FIG. 2B by one of ordinary skill in the art, the intersection of the axis labeled spring potential energy, the axis labeled displacement, and the curve 252 represents a state in which the valve 100 tends to return after being disturbed and which point is commonly understood as stable equilibrium. Accordingly, the middle position 212 at the bottom of the potential energy well is a stable equilibrium point. The fact that the ball 250 is on a steep incline as shown is consistent with the large restoring force being applied to the valve 100 in the open position. The large restoring force of the spring 104 and 106 implies that the linear motor (i.e., electromagnets 108 and 110 together with the armature 112) must apply large holding forces to maintain the valve 100 in the open or closed position which is often undesirable as this requires electrical power.
In light of the above, a more desirable ideal shape for a potential energy function is that shown by the dashed curve 254 in FIG. 2B and corresponding ideal nonlinear restoring force curve 202 in FIG. 2A. Potential energy curve 254 reaches a peak at the open and closed positions corresponding to zero force in FIG. 2A where correspondences are indicated with dashed lines 206. If such an ideal nonlinear spring were created, the holding forces of the linear motor of FIG. 1 would be minimal and the associated jerk in the valve motion would also be minimized. In order to operate a valve with total spring potential energy of the ideal form shown as 254, such a linear motor would have to apply a small push to move the valve off the peaks of potential energy during valve transitions.
It is also known that the ideal nonlinear restoring force curve 202 in FIG. 2A can be, for example, accomplished with a cam and spring. The present discussion refers to any mechanism where the magnitude of the force or torque required to hold the valve open or closed is less than a peak spring force as a “nonlinear spring.” Nonlinear springs are usually designed using cams in combination with springs generally abiding by Hooke's law. The ideal nonlinear restoring force curve 202 has a peak value which occurs at 204 and there is an intersection 210 where the curves 204 and 202 cross the displacement axis (with positive slope) and a corresponding zero total spring restoring force at the crossing point. Such intersection 210 is the point at which the valve 100 tends to return after being disturbed and which point is commonly understood as stable equilibrium. Accordingly, the intersection 210 is a stable equilibrium point. Whether the spring characteristic is linear or nonlinear, it is readily apparent that the given linear motor must supply any energy lost to friction, impacts, and vibration.
One example of a nonlinear spring can be realized with the prior art system of FIG. 3A. Cam 300 works in conjunction with cam followers 302 and 304, and two springs 306 and 308 to create the potential energy characteristic of FIG. 3B which has an associated nonlinear spring curve (not shown) similar to 202. Note that the ordinate in FIG. 3B is the motor displacement angle θ. The motor 310 is coupled to the cam 300 in any known manner (e.g., by gears, belts, or direct co-axial drive) as generally indicated by the dashed line 312. The motor executes a reciprocating motion and causes a reciprocating rotary motion of +/−45° in the cam relative to the θ=0° reference angle 314. The cam 300 is shown in the +45° position. An important property of this design is the fact that a positive follower force is applied consistently to the cam 300 during operation as the springs 306 and 308 are always under some level of compression. The present disclosure refers to such a follower and cam combination a “lashless cam.” In contrast, a follower and cam combination whose contact force becomes less than or equal to zero at some point during its operation is called a “lashing cam.” Lashing cams have the undesirable property of having impact between the follower and the cam which cause wear, vibration and energy loss.
It should be noted that the positive contact force in FIG. 3A is applied intrinsically by the energy storing springs 306 and 308 rather than extrinsically with a separate preload mechanism whose primary function is to apply force or absorb lash in the system. The present disclosure refers to any follower and cam combination held in positive contact using an energy storage spring as an “intrinsic lashless cam.” Such cam and follower combinations are desirable due to their simplicity in addition to the reduction in impact, vibration, and noise associated with the elimination of lash in the valve system.
FIG. 3B shows the curve 350 for the total potential energy in the springs 306 and 308 for the system of FIG. 3A. The symmetry of the curve follows from the symmetry of the cam 300, followers 302 and 304, and identical maximum compression and spring constants of the springs 306 and 308. All of these factors can be varied to shape the total potential energy curve. For example, one prior art design aims to minimize the peak energy 352 in order to minimize the torque supplied by the motor. Further, the 90° angle shown between the springs 306 and 308 can be varied to shape the total potential energy curve. In addition to the compressive force that the cam 300 applies to the followers 302 and 304 and springs 306 and 308, the cam 300, at some points in the cycle, also disadvantageously applies a tangential force to the followers 302 and 304 and a bending force the valve stem 315 and pin 316, induces bending vibrations in the valve stem 315 and pin 316, and increases frictional forces in the valve guide 318 and pin guide 320. Because the valve guide 318 often serves as a gas seal, it is difficult to mitigate frictional losses with rolling element bearings or other such mechanisms in the valve guide 318. In the present disclosure, this type of cam follower producing tangential or bending forces is described as a “non-zero tangential force cam” and those that do not produce tangential or bending forces are described as a “zero tangential force cam.” It should be understood though that exactly zero tangential force is not attainable due to manufacturing and other imperfections.
FIGS. 4A, 4B, and 4C depict another prior art electronic valve system. FIG. 4B is a section of the disk cam 400 in FIG. 4A along section 4B-4B. FIG. 4C shows the total potential energy function 460 for springs 402 as a function of motor angle θ in FIG. 4A.
FIG. 4A shows the half-open position for valve 404 corresponding to the minimum total potential energy of the springs 402. Motor 406 drives the disk cam 400 in a rotary oscillatory fashion. FIG. 4B shows section 4B-4B of FIG. 4A. The cam follower 408, typically a roller follower, follows the cam slot profile 430 during valve transitions. The cam slot profile surface is nearly tangent to a concentric circle at 432 in the valve open position, and nearly tangent to a second concentric circle at 434. The axis of cam rotation in FIG. 4B corresponds to the center of the concentric circles 432, 434. This near tangency condition at 432 and 434 implies that the motor torque required to hold the valve closed or fully open is advantageously nearly zero. The flats 462 and 464 in FIG. 4C correspond to the mentioned tangencies in the cam profile.
With further regard to FIG. 4A, as one of the equilibrium positions of the springs 402 is the half-open position, the follower 408 applies an upward force along the positive Z direction (shown by arrow at bottom of FIG. 4A) when the valve 404 is closed and a downward force along the negative Z direction when the valve 404 is fully open. As a consequence, the cam 400 and follower 408 combination is a lashing cam. The follower must alternately contact the inner and outer profile surfaces of the cam slot 430. This lash disadvantageously causes impact, vibration, and energy losses in the valve system. In addition, the force 436 applied to the follower 408 by the cam 400 has a non-zero tangential component 415. Thus, such cam system is a non-zero tangential force cam. Accordingly, it applies a bending force to the valve stem 410 thereby creating bending vibrations of the valve stem 410, and contributes to friction in the valve guides 411. Another drawback of this design is the valve seat positioning problem associated with maintaining tolerance on dimension 412 between the seat and the cam profile surface at 434.
FIGS. 5A, 5B, and 5C show another known variation of that shown in FIG. 4A which can overcome the valve seat positioning problem because the cam profile is parallel to the valve motion in the seated position. FIG. 5B shows section 4B-4B of FIG. 5A. Rotary spring 500 together with disk cam 502 and cam follower 504 create the spring potential energy curve 558 shown in FIG. 5C. Note that the ordinate in FIG. 5C is the vertical displacement Z of the valve and linear motor 506. The minimum of curve 558 corresponds to the valve half-open position. The cam profile 530 is tangent to radial lines 532 and 534 at the ends of the motion so that the spring potential energy has flats 560 and 562 respectively. Consequently, the motor force opposing the spring force when the valve is closed or fully open is nearly zero. The linear motor must counter any gas forces on the valve in contrast to the design of FIG. 4A. Near the ends of the valve motion, the spring displacement forces are high. Hence, there is a large cam follower force 536 shown in FIG. 5B and a large tangential force 538. Thus, this prior art design provides a non-zero tangential force follower with the associated problems of bending forces and vibration and increased friction in the guides 508. Finally, this is a lashing cam system as the cam torque switches sign at the half-open point. Thus, there is impact, vibration, and energy loss as the cam follower moves from one surface of the cam slot to the other.
Thus, it is desirable to create a mechanical valve mechanism with a natural motion that is nearly the desired opening/closing motion of the valve and to augment this mechanism with an electric motor to control timing and compensate for losses due to friction, impacts, mechanism lash, vibration and other loss occurrences. It is desired, in addition, to have such an electrically actuated valve mechanism that has minimal friction, minimal bending forces on the valve stem, minimal jerk in the valve motion, minimal unwanted vibration, and minimal impacts to due to lash in the mechanism. It is also desired that the valve mechanism have minimal mass because the motor is needed to initiate the opening/closing of the valve and to stabilize the valve motion upon closing. Finally, it is desirable to have motor current waveforms that achieve the desired opening and closing times of the valve and also minimize the electrical power consumed by the electric motor.